The present invention relates to a variable displacement compressor and, more particularly, to a variable displacement compressor having axial pistons carried by a swash plate the tilting angle of which is varied to control the stroke of the pistons and, hence, the displacement of the compressor.
In general, a variable displacement compressor of the type mentioned above has a swash plate mechanism adapted to be rotated by a shaft which in turn is driven by the power of, for example, an engine. The swash plate mechanism carries a plurality of axial pistons. The angle of tilt of the swash plate mecahnism is changed so as to vary the stroke of the axial pistons, thereby controlling the displacement of the compressor. More specifically, the control of the displacement is conducted by varying the angle of tilt of the swash plate about its fulcrum or supporting plate. Various types of mechanism are available for effecting this control.
For instance, the specification of U.S. Pat. No. 2,964,234 discloses a mechanism in which, as shown in FIG. 4 of the present application, a torque-transmitting lug 41 provided on a drive shaft 40 carries a falcrum pin 44 which engages with a cam groove 43 formed in a swash plate 42. In operation, a sleeve 45 which rotatably supports the swash plate 42 is displaced axially and the cam groove is made to move relative to the fulcrum pin 44, thereby effecting tilting action and the control of tilting angle of the swash plate 42. This arrangement is characterized in that the fulcrum pin 44 is kept stationary while the cam groove 43 is moved.
Another known mechanism disclosed in the specification of U.S. Pat. No. 4,061,443 has, as shown in FIG. 6, a torque-transmitting lug 51 provided on a drive shaft 50 has a cam groove 52 which engages with a fulcrum pin 54 on a swash plate 53. The swash plate 53 is connected through a connector pin 57 to a sleeve 55 slidably mounted on the drive shaft 50. In operation, the sleeve 55 is made to slide on the drive shaft 50 and the fulcrum pin 54 is made to move along the cam groove 52 in the lug 51, thereby conducting the tilting operation and the control of the tilting angle of the swash plate 53. Thus, this arrangement features a combination of a movable fulcrum pin 54 and a stationary cam groove 52, in contrast to the known arrangement explained in connection with FIG. 4.
A further known arrangement is proposed in the specification of U.S. Pat. No. 4,178,135. As shown in FIG. 9, this arrangement has a drive plate 61 carried by a drive shaft 60 for rotation as a unit therewith. A cam surface 62 is formed on one axial end face of the drive plate 61. A fulcrum roller 64 carried by a swash plate 63 is urged by a spring 65 into contact with the cam surface 62. The tilting operation and the control of the tilting angle of the swash plate 63 are effected by a movement of the fulcrum roller 64 along the cam surface 62 and by an axial movement of a support pin 66 which supports the swash plate 63. This known arrangement features a combination of the movable fulcrum roller 64 and the stationary cam surface 62.
These known arrangements, however, suffer from the following disadvantages:
In the first-mentioned known arrangement shown in FIG. 4, the position of the fulcrum, i.e., the position of the fulcrum pin 44, is not changed even though the tilting angle .theta. of the swash plate 42 is changed, because this arrangement employs a combination of the stationary fulcrum pin 44 and the movable cam groove 43 in the swash plate 42. In consequence, as shown in FIGS. 5A and 5B, the distance y.sub.p between the point of action of the reactional force F.sub.G of the compressed gas transmitted from the pistons and the fulcrum point is constant, so that, if the reactional force F.sub.G of the compressed gas is constant, the tilting moment M.sub.T (M.sub.T =F.sub.G .multidot.y.sub.p) of the swash plate 42 about the fulcrum point is constant. The control of the tilting angle of the swash plate 42 relies upon the balance between the tilting moment M.sub.T produced by the reactional force F.sub.G produced by the compressed gas and either a counter control force F.sub.C which is produced by the pressure of the gas in the crank chamber or a load applied to the pin 45 on the swash plate 42. In order to reduce the tilting angle of the swash plate so as to reduce the piston stroke, it is necessary to increase either the control force F.sub.C produced by the internal pressure of the crank chamber or the load applied to the pin 45. For attaining a good response to the control input for controlling the tilting angle, it is desirable that the control force F.sub.C is made as small as possible. In other words, a higher tilting angle response characteristic can be obtained by decreasing the increment of the control force F.sub.C required for the control. Since the tilting angle is maintained by the balance of force between the tilting moment M.sub.T and the control force F.sub.C, the reduction in the increment of the control force F.sub.C essentially requires that the tilting moment M.sub.T is decreased. To this end, it is necessary that the distance y.sub.p between the fulcrum of the swash plate and the point of action of the reactional force F.sub.G of the gas is decreased as the tilting angle of the swash plate decreases. To this end, a mechanism is required which would progressively move the fulcrum point 44 towards the drive shaft 40 as the tilting angle decreases. In particular, where the discharge pressure and the suction pressure are constant, the distance y.sub.G between the point of action of the reactional force of the compressed gas and the center of the drive shaft is maintained constant, so that a reduction in the distance y.sub.p is essentially required for the purpose of reducing the tilting moment M.sub.T and, hence, the control force F.sub.C. Unfortunately, however, the known arrangement shown in FIG. 4 fails to meet this requirement because the point of the fulcrum, i.e., the point of the pin 44, is unchangeable with respect to the point of action of the reactional force F.sub.G.
Referring now to the arrangement explained in connection with FIG. 6, the problem encountered by the first-mentioned known arrangement is overcome because the fulcrum pin 54 provided on the swash plate 53 is movable while the cam groove 52 on the side of the drive shaft 50 is stationary. In this arrangement, the lug 51 is projected through a slot 56 formed in the sleeve 55 which is slidably carried by the drive shaft 50. This construction has the following problem: Namely, as will be seen from FIGS. 7 and 8, the effective area indicated by A with marks "x" for receiving the pressure between the sleeve 55 and the drive shaft at the half circumference of the shaft where the slot 56 is formed, is smaller than at the other half circumference of the shaft. In consequence, a radial component P of the reactional force F.sub.G of the compressed gas transmitted by the pistons acts more heavily on this area A than on the other circumference, resulting in local wear of the sleeve 55 and the drive shaft 50 during repetitional sliding movements of the sleeve 55 on the drive shaft 50.
Referring now to the third known arrangement shown in FIG. 9, the fulcrum roller 64 constituting the fulcrum point for the tilting motion of the swash plate 63 is prevented from moving apart from the cam surface 62 on the drive plate 61 by virtue of the spring 65 which acts to urge the fulcrum roller 64 into contact with the cam surface 62. When the compressor is started, however, a force which tends to move the fulcrum roller 64 from the cam surface 62 is generated by the inertia of the pistons 67 and the friction between the pistons 67 and the respective cylinder bores 68. In consequence, the fulcrum roller 64 is momentarily separated from the cam surface 62 and is then sprung back into contact with the cam surface 62 by the force of the spring 65 and the force produced by the compressed gas. In consequence, the cam surface 62 and/or the fulcrum roller 64 is heavily worn due to repeated collisions therebetween. The cam contour of the cam surface 62 and the shape and size of the support roller 64 must be strictly managed so as to provide a constant top clearance between the free axial end of each piston and the top wall of the associated cylinder bore. However, the heavy wear of the support roller 64 or the cam surface 62 undesirably increases the top clearance.